This invention relates generally to free-piston Stirling machines and more particularly relates to non-contact bearing support systems that support their power piston and/or displacer piston and their respective connecting rods attached to them. The invention improves the life, reliability and cost of free-piston machinery by providing a simple and reliable means to implement non-contact bearings in a manner that reduces the difficulty of aligning the bearings or allows more accurate alignment or both.
Although free-piston Stirling cycle machines have been shown in the prior art in a very extensive variety of configurations, most have a displacer piston and a power piston that reciprocate in the same cylinder or in different cylinders. An end of the power piston and often an end of the displacer piston is ordinarily rigidly fixed to a connecting rod that reciprocates with the piston. These components together as a unit are supported within a casing of the Stirling machine. The casing contains a working gas that alternately expands and compresses as the working gas is shuttled between an expansion space and a compression space.
Stirling machines are designed to provide either: (1) an engine having a power piston and displacer piston driven by applying an external source of heat energy to the expansion space and transferring heat away from the compression space and therefore capable of being a prime mover for a mechanical load, or (2) a heat pump having the power piston (and sometimes the displacer piston) cyclically driven by a prime mover for pumping heat from the expansion space to the compression space and therefore capable of pumping heat energy from a cooler mass to a warmer mass. The heat pump mode permits Stirling machines to be used for cooling an object in thermal connection to its expansion space, including to cryogenic temperatures, or heating an object, such as a home heating heat exchanger, in thermal connection to its compression space. Therefore, the term Stirling “machine” is used to generically include both Stirling engines and Stirling heat pumps, the latter sometimes being referred to as coolers. Both Stirling engines and Stirling heat pumps, like electromagnetic motors and generators or alternators, are both basically the same power transducer structures capable of transducing power in either direction between two types of power.
In order to minimize the frictional wear of the reciprocating components of a free-piston machine, it is desirable to avoid contact between the reciprocating bodies and their cylinders or other supports within the casing. Conventional lubricants cannot be used for this purpose because they substantially degrade the properties of the working gas and result in a substantial decrease in the efficiency of the free-piston Stirling machine. For these reasons, free-piston Stirling cycle machines commonly use gas bearings and also radially acting spring bearings, such as planar springs. Although both kinds of bearings are known in the art, some explanation of gas bearings and radially acting spring bearings is desirable because some aspects of their operation are relevant to the invention.
A bearing is a device that supports, guides, and reduces the friction of motion between at least two parts that move with respect to each other. A bearing supports the two parts in a relative position or orientation with respect to each other but permits one part to move with respect to the second part in one or more directions of motion. It is often desirable to minimize the friction between the parts and minimize the force applied by one part to the other in the permitted directions of motion. A “non-contact bearing” supports the parts in a manner that the parts themselves that are moving relative to each other do not come into contact. The bearing itself, such as a planar spring bearing, may contact both parts, but it does not rub or slide against either part.
A gas bearing is one type of non-contact bearing that is often used on free-piston Stirling machines to maintain the separation of a piston in a cylinder or a connecting rod in a cylindrical bore. The gas bearing uses a gas, typically the working gas, that is pumped between relatively moving surfaces and functions as a lubricant to maintain separation of the relatively moving surfaces. Gas bearing systems have a fluid flow loop in which working gas is pumped out of ports in the piston or cylinder into the clearance gap between the piston and cylinder. To construct an effective gas bearing, the clearance fit between the two moving surfaces must be a close fitting clearance and the distance range of that clearance for a gas bearing in a Stirling machine is known to those skilled in the art. There must be at least three such ports spaced around the cylindrical periphery, preferably equi-angularly (every)120°, so that there will be radially inwardly directed centering forces applied toward centering the piston regardless of the radial direction in which the piston may become off center. Because gas bearings require close fitting clearances, if a cylindrical surface of one body has a close fit clearance with a cylindrical surface of another body because there is a gas bearing between them, the axes of the two cylindrical surfaces must be aligned to avoid contact.
A close fit clearance between a cylindrical surface of one body with a cylindrical surface of another body can also provide a “clearance seal”. It is commonly desirable to provide a seal between two parts, such as a piston and the associated cylinder in which it reciprocates. The seal is intended to prevent or minimize the flow of a fluid between the piston and cylinder from one end of the piston to the other. However, it is desirable to simultaneously prevent contact between the piston and its cylinder in order to prevent wear and therefore gas bearings are used. Although not perfect, the clearance between the piston and its cylinder can be made sufficiently small to provide both reasonably effective sealing as well as a non-contact bearing. Such a seal using a small clearance fit is a clearance seal. The “seal length” of a clearance seal may be defined as the effective length in the axial direction of the portion of the piston's cylindrical periphery that is formed as the clearance seal; that is, the close fit clearance portion. Most commonly, that is the entire length of the piston. However, if the piston at times is displaced along the cylinder to a position where it protrudes from the cylinder, then the effective seal length of the clearance seal is shortened slightly and more particularly is the time averaged length of the clearance seal interface between the piston and its associated cylinder. The “axial center” of the clearance seal may be defined as the center, along the axial direction, midway between the axially opposite ends of the clearance seal. That midway position is the axial center and can be used to define the position of the clearance seal.
A radially acting spring bearing is another type of non-contact bearing that has been used on free-piston Stirling machines. Although the term “radially acting spring bearing” is not commonly used, it has been adopted because it is believed to best describe one of the bearings that is used in embodiments of the invention. A “radially acting spring bearing” is a spring that is connected to each of the two bodies that are to be supported in a non-contact relationship with one body moving with respect to the other. This bearing applies its spring force in a radial direction opposite its radial direction of deflection from its central axis when it is deflected away from its relaxed condition at the central axis. Its spring force in a radial direction is 0 for no deflection from its axis which means that it introduces no side loading. It can additionally apply a spring force in an axial direction so that it has two components of spring force, axial and radial. So a radially acting spring bearing is a spring that has a component of force in the radial direction, applies no radial force when centered and its force in the axial direction can be 0 or finite. For the invention, it should apply no significant net side forces as it is deflected.
An example of a commonly used radially acting spring bearing that is known in the prior art is a planar spring. A planar spring typically has arms extending from a central hub to an outer rim along a spiral-like or involute-like path. The arms, hub and rim are usually in a plane in their relaxed state. Typically the arms have a width in the plane considerably greater than their thickness perpendicular to the plane. Planar springs used as bearings are very stiff for deflection in the radial direction, but also apply a spring force, with far less stiffness, when deflected in the axial direction.
A common coil spring, in which a wire is wound as a helix, cannot be used as a radially acting spring bearing if oriented in an axial direction because it applies significant side forces when deflected axially. However, it would be possible to use several radially oriented coil springs arranged along radials of an axis of reciprocation as a radially acting spring bearing. Also usable is a spiral or involute spring, similar to a planar spring and typically constructed of spring wire wound in a plane along a spiral-like pattern, with connections to the other machine components at the innermost, centrally located end of the wire and at the outermost peripheral part of the wire. A conical coil spring might also be used but risks the introduction of side loads like the coil spring.
Great effort has been expended in the prior art in order to avoid oil-type lubricants to prevent wear of the internal components of Stirling cycle engines and coolers while avoiding contamination of the working gas. The free-piston configuration greatly reduces side loads because the free-piston configuration does not use a motion translating mechanism that introduces side loads, such as a connecting rod connected to a crankshaft. However, it is still necessary to provide bearing support for a reciprocating part in order to avoid excessive wear. Two techniques in the prior art have found common application to solve the problem of supporting a free piston that has a close fit clearance in a manner that avoids contact between the close fit surfaces and yet allows reciprocation of the piston.
The first technique, referred to as flexural bearing support (e.g. U.S. Pat. No. 5,920,133, Penswick et al and U.S. Pat. No. 5,522,214, Beckett et al), is to support the moving components entirely on planar springs so that there is no contact between the cylinder and the moving component (power piston or displacer piston). This bearing support system is shown in FIG. 1 implemented on a posted-displacer configuration free-piston Stirling machine. A piston 2 is supported by flexures 4 and 6 at points 8 and 10 on the piston 2 so that close-fitting clearance A is maintained with cylinder 12. The displacer 14 is similarly supported by flexures 16 and 18 at points 20 and 22 so that close-fitting clearances B and C are maintained. All of these flexures are planar springs. Flexures 4 and 6 are securely held on support structure 24 so that there is essentially no radial motion while providing limited axial motion. The support structure 24 is fixed to the casing 26 so that the peripheral rim portion of the flexures 4 and 6 are effectively fixed to the casing 26. “Fixed to the casing” means attached directly or indirectly in a fixed position relative to the casing because a component part can be fixed to an interposed structure that is itself fixed to the casing. Flexures 16 and 18 are supported peripherally on the displacer 14 and at their centers on the displacer rod 28. The displacer rod 28 is rigidly attached to the cylinder 12 which in turn is fixed to the casing 26. A linear alternator/motor 30 provides electrical output or mechanical input depending on whether the free-piston machine is an engine or a heat pump, respectively. The casing 26 is hermetically sealed and contains the moving parts.
The problem with the prior art of FIG. 1 is that the flexures 4 and 6 must be precisely aligned so that the power piston 2 is unable to make contact with the cylinder 12. Similarly, the flexures 16 and 18 must be precisely aligned so that the displacer piston 14 is unable to make contact with the cylinder 12. Furthermore, the flexures must be sufficiently stiff to support the piston weight if the machine runs with a non-vertical axis of reciprocation in a gravitational field and to support the pistons against other side loads.
The difficulty of this problem of alignment is illustrated in FIG. 2 which is a diagram showing a piston 40 that reciprocates in a cylinder 42. The clearance is greatly exaggerated in order to illustrate the applicable principles. The piston 40 has a connecting rod 44 fixed coaxially to an end of the piston. As used in this description, a “connecting rod” is an essentially rigid link connecting a piston to another component. Commonly, a connecting “rod” is a solid cylindrical rod but it is not necessary that the connecting rod be a solid material throughout its cross section and it is not necessary that it have a cylindrical peripheral surface or even a symmetrical outer peripheral surface when viewed in cross section. For example a connecting rod can be a tube and or have an I-beam or L-beam cross-section. Therefore the term “rod” is used but is not limited to a solid rod but includes other shapes of rigid connecting arms, including multiple smaller arms that together act mechanically as a single connecting arm. Ordinarily, the connecting rod is connected to an axially reciprocating load that is driven by the Stirling machine or a prime mover that drives the Stirling machine. Since it is desirable to minimize the volume of a machine, the “connecting rod” of a power piston can have components of the load or prime mover mounted to it in such a manner that a separate connecting rod is not readily apparent. That is the case with the structure of FIG. 1 in which the reciprocating magnets 54 and 56 of the linear alternator or motor are mounted to a connecting rod that has the same diameter as the piston 2 and is not visibly distinguishable from the piston, although it is functionally distinguishable. Furthermore, the “connecting rod” of FIG. 1 also connects the piston to two flexures 4 and 6 and has a component of the linear alternator/motor interposed between its ends. All these characteristics can be characteristics of a connecting rod.
As seen in FIG. 2, the proper alignment of the piston 40 in the cylinder 42 requires that two points, 46 and 48, be accurately positioned. One point is the intersection of the axis of the piston and a plane perpendicular to the axis at one end of the piston (or more concisely at one end of the close fit clearance). The second point is the intersection of the axis of the piston and a plane perpendicular to the axis at the opposite end of the piston (or more concisely at the opposite end of the close fit clearance). The rightmost two black dots in FIG. 1 illustrate the corresponding points for the embodiment of FIG. 1. Those two intersection points must both be positioned on or very near the axis 49 of the cylinder 42 in order to avoid contact of the outer periphery of the piston with the surface of its cylinder. However, as illustrated in FIG. 2, any rotation of the piston 40 and its connecting rod 44 away from coaxial alignment also moves the axis 51 of the connecting rod 44 radially away from the axis 49 of the cylinder 42. At some sufficient angle of misalignment, the peripheral surface at one or both ends of the piston 40 will contact the cylinder 42 as illustrated by dashed lines.
Referring again to FIG. 1, an extension of the piston 2 protrudes out of the cylinder 12 and into the reciprocating component of the electric linear motor or alternator. That extension functions as a connecting rod which couples the motion of the piston 2 of the Stirling machine to the linear motor/alternator. Because that connecting rod is displaced off-center by any misalignment of the piston, in the structure of FIG. 1, it is necessary to simultaneously align two additional points 50 and 52 along the axis of the cylinder 12. Those two additional points 50 and 52 are the intersection 50 of the axis of the piston 2 with a plane perpendicular to that axis at the attachment point of the flexure 4 to the piston 2 and the intersection 52 of the axis of the piston 2 with a plane perpendicular to that axis at the attachment point of the flexure 6 to the piston 2. The problem solved by the invention arises because of the difficulty of obtaining accurate alignment of four points symbolized by the four black dots in FIG. 1. The problem is that radial adjustment of any one point moves the radial position of at least two of the three other points. Of course only the positions of the two flexures 4 and 6 can be manipulated in the alignment procedure. But the movement of one always affects the required position of the other. So the adjustment procedure always requires going back and forth between the two flexure adjustments and is difficult and time consuming to accomplish satisfactory alignment.
FIG. 3 illustrates a beta free-piston Stirling machine with gas bearings, indicated by radially inwardly directed arrows, and with a planar spring 60 as a bearing. A displacer piston 62 reciprocates in a cylinder 64 and has a close fit clearance 66 that is needed for its gas bearing. A power piston 68 reciprocates in the cylinder 64 and is separated from it by a gas bearing formed at the close fit clearance 70. A connecting rod 72 is fixed at one end to the end of the displacer piston 62 and at its opposite end to a planar spring bearing 60. The connecting rod 72 has a cylindrical exterior and extends through a cylindrical bore axially through the piston 68. A gas bearing is formed at the close fit clearance 74 between the connecting rod 72 and the piston 68.
For the displacer piston 62 and its connecting rod 72, there are five points that must be aligned and they are illustrated by the large black dots, not including point 75. There are two points for the gas bearing at the close fit clearance 66, for the reasons explained above, two points for the gas bearing at the close fit clearance 74 and one point for the planar spring bearing 60. For the piston 68 there are five points that must be aligned not including point 77, two for the gas bearing at the close fit clearance 74, two for the gas bearing at the close fit clearance 70 and one point for the planar spring bearing 60.
In order to alleviate the problem of aligning five points, the prior art discloses an implementation of gas bearings with compliance built into the connecting rod as illustrated in FIG. 4 for a beta free-piston Stirling machine. A piston 80 is supported by gas bearings at close-fitting clearance 82 between the piston 80 and the cylinder 84. A displacer piston 86 is similarly supported in the cylinder 84 by gas bearings at the close fitting clearance 88. A connecting rod 90 is connected to the end of the displacer piston 86 and is supported by a gas bearing at close-fitting clearance 92 along the interfacing exterior of the connecting rod 90 and the interior of the axial bore through the piston 80. In order to avoid excessive side-loads and/or assembly tolerance stack-up, the planar spring 94 is connected to the displacer rod 90 by way of flexure rod 96 which is a compliant member. As in the devices of FIGS. 1 and 3, a linear alternator/motor 98 provides electrical output or mechanical input depending on whether the machine is an engine or a heat pump.
As in FIG. 3, the power piston 80 is supported on gas bearings at its peripheral, cylindrical surface, the displacer piston 86 is supported on gas bearings at it peripheral surface and on the displacer connecting rod 90 where the connecting rod 90 is within the piston 80. The compliant member 96 is used to connect the displacer rod 90 to the planar spring bearing 94. The planar spring 94 may provide additional radial compliance to reduce side loads on the displacer due to constructional inaccuracies. The basic concept of using a compliant flexure rod 96 to connect the end of the connecting rod 90 to the planar spring bearing is that the point of the attachment of the compliant flexure rod 96 to the planar spring bearing is not as critical because the machine can operate with the compliant flexure rod 96 in a slightly bent condition without introducing excessive side loading. Therefore, less accurate positioning of that attachment point can be tolerated. Nonetheless, there remain four points that must be aligned as illustrated by the black dots on FIG. 4.
The chief difficulty of this arrangement is that in order to obtain satisfactory stiffness on the displacer rod gas bearing, a very close fit of less than 25 μm diametrical clearance is required with the bore in the piston. In some cases, particularly smaller machines where the rod may be only around 3 to 5 mm in diameter, the clearance may be as small as 8 to 15 μm. This places a requirement of precision that cascades through the structure resulting in further precision requirements of concentricity, straightness and perpendicularity.
The flexural system of FIG. 1 is highly limited in amplitude and requires substantial space for implementation and is therefore associated with bulky configurations. The planar spring bearings 4, 6, 18 and 20 must be sufficiently stiff to support the piston weight if the machine runs on its side in a gravitational field (i.e. with its axis not vertical) and other side loads. Furthermore, since the planar springs are responsible for holding the clearance between the moving part and its cylinder, an extraordinary level of precision is required for the components and their assembly. The conventional gas bearing technique of FIG. 4 has a more relaxed precision but suffers from very feeble support on small diameters, e.g., the displacer rod on free-piston Stirling machines. Thus, a requirement of this technique is to employ compliance so that other components attached to the moving parts (mechanical springs, for example) will not overcome the gas bearing load capacity (e.g., U.S. Pat. No. 5,525,845, Beale et al).
The above description demonstrates that the bearing systems that have been shown in the prior art require a high degree of precision in the machining of parts and a high degree of precision in the alignment of parts or are limited by very feeble support of gas bearings on small diameters. The purpose of the invention is to reduce the degree of precision required for alignment while maintaining the other favorable characteristics of non-contact bearings.
An ideal bearing system for piston-cylinder assemblies, particularly for use in free-piston machinery, would have the following attributes in addition to non-contact operation:                a. No greater precision required than that for satisfactory performance from the machine. That is, the bearing system should minimize the requirement of additional precision components.        b. The bearing system should require no end-loop adjustments during manufacture.        c. The bearing system should be robust so that there is no possibility of the bearings going out of adjustment over time.        d. The bearing system should be able to tolerate a reasonable level of external shock or component over stroke without becoming misaligned.        
The proposed invention has these advantages over current systems.